EFFECT OF A TIP CLEARANCE ON THE … design details of the impeller which is ... Blade angle at inducer ... dimensional computational mesh generation process for compressor blade with

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International Journal of Mechanical Engineering and Technology (IJMET) Volume 8, Issue 1, January 2017, pp. 178–188, Article ID: IJMET_08_01_020

Available online at http://www.iaeme.com/IJMET/issues.asp?JType=IJMET&VType=8&IType=1

ISSN Print: 0976-6340 and ISSN Online: 0976-6359

© IAEME Publication

EFFECT OF A TIP CLEARANCE ON THE

PERFORMANCE OF A LOW SPEED CENTRIFUGAL

COMPRESSOR

S.M. Swamy

Assistant Professor, Department of Mechanical Engineering,

GNITS, Shaikpet, Hyderabad, India

V. Panndurangadu

Professor, Department of Mechanical Engineering,

JNTUA, Ananthapuramu, A.P, India

J.M. Shamkumar

Design Engineer, Airframe Aero designs Pvt. Ltd., Bangaluru, India

ABSTRACT

Tip clearance effects on flow field of a low speed centrifugal compressor without and with

partial shroud (PS) attached to the rotor blade tip at three values of tip clearance, viz. φ =2.2%,

5.1% and 7.9% of rotor blade height at the exit at three flow coefficients, namely, φ=0.12,0.18

(below design flow coefficient), φ=0.28 (design flow coefficient) and φ=0.34 (above design flow

coefficient), is analysed computationally using structured multi block grid with fine grid in the tip

clearance region. The paper aims to study several flow characteristics between blade channels

using commercial flow solver ANSYS CFX 15.0 based on finite volume techniques. The centrifugal

compressor in aerodynamic requirement is that edge velocities along the impeller channel passage

surfaces like hub, shroud, pressure and suction surfaces vary smoothly without sudden

decelerations, which cause flow separation leading to losses. Using the periodic boundaries and

defined flow conditions at inflow / exit flow and blade rotations, the turbulent viscous flow between

blade channels are computed. The efficiency related parameters using average quantities, besides

flow pattern in terms of velocities, streamlines and pressure distribution on blade surfaces are

graphically interpreted. An attempt is also made to study the influence of pressure loads on

structural deformations in the chosen blade profile. This paper highlights aero-mechanical features

of centrifugal impeller obtained from several numerical simulations, which are expected to provide

a sound basis for further investigations.

Key words: Centrifugal compressor, Flow coefficient, Partial shroud attached to the rotor blade tip,

Streamlines, passage wake.

Cite this Article: S.M. Swamy, V. Panndurangadu and J.M. Shamkumar. Effect of a Tip Clearance

on the Performance of a Low Speed Centrifugal Compressor. International Journal of Mechanical

Engineering and Technology, 8(1), 2017, pp. 178–188.

http://www.iaeme.com/IJMET/issues.asp?JType=IJMET&VType=8&IType=1

Effect of a Tip Clearance on the Performance of a Low Speed Centrifugal Compressor

http://www.iaeme.com/IJMET/index.

1. INTRODUCTION

The centrifugal compressors have a wide range of applications especially for power plants for small

aircraft and helicopters, in process industries, compression of gases and vapours, because they can provide

high-pressure ratios and large operating ranges

are used primarily for their suitability for handling small volume flows, but other advantages include a

shorter length than an equivalent axial flow compressor, less susceptibility to loss of perf

buildup of deposits on the blade surfaces and their suitability to operate over a wide range of mass flow.

The gas turbine engines powering the most passenger and military aircrafts are also called jet engines.

clearance in centrifugal compressor causes the leakage of high pressure fluid from pressure surface to

suction surface of the impeller blade, making the flow field highly complex and effecting the performance.

In centrifugal compressor due to the low efficiency and thrust

range applications. However, the inability for spark ignition engines to work at high altitude warrants

further development to improve the efficiency of these miniature gas turbines.

from his experiments that axial movement of the casing has better efficiency over the movement of casing

in radial and axial directions. Radial movement of casing increases clearance at inducer, which reduces the

operating range. The tipclearance studies are conducted to und

minimise the effect of tip clearance. Pampreen (1973), Mashimo et al. (1979), Sitaram and Pandey (1990)

have conducted experimental studies and suggested that by reducing the tip clearance gap size, the tip

clearance effect can be minimised.

2. MODELING OF COMPRESS

The design details of the impeller which is used in the present investigation are given below:

Inducer hub diameter, d1h

= 160 mm

Impeller tip diameter, d2

= 500 mm Blade height at the exit, b

No. of blades of impeller, Nb

= 16

Blade angle at inducer tip, β1t

= 35

Thickness of the blade, t = 3 mm

All angles are with respect to the tangential direction.

Figure 1 Schematic diagram of centrifugal compressor setup


IJMET/index.asp 179



pressure ratios and large operating ranges with relatively high efficiencies. Centrifugal compressors


shorter length than an equivalent axial flow compressor, less susceptibility to loss of perf


The gas turbine engines powering the most passenger and military aircrafts are also called jet engines.

essor causes the leakage of high pressure fluid from pressure surface to


In centrifugal compressor due to the low efficiency and thrust to weight ratio, they are limited to short

However, the inability for spark ignition engines to work at high altitude warrants

further development to improve the efficiency of these miniature gas turbines.

s that axial movement of the casing has better efficiency over the movement of casing


operating range. The tipclearance studies are conducted to understand the flow behavior in order to



ffect can be minimised.

MODELING OF COMPRESSOR BLADES


= 160 mm Inducer tip diameter, d1t

= 300 mm

= 500 mm Blade height at the exit, b2

= 34.7 mm

= 16 Blade angle at inducer hub, β1h

= 53o

= 35o

Blade angle at exit, β2

= 90o

Rotor speed, N = 2000 rpm

All angles are with respect to the tangential direction.

Schematic diagram of centrifugal compressor setup and blade to blade view


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with relatively high efficiencies. Centrifugal compressors


shorter length than an equivalent axial flow compressor, less susceptibility to loss of performance by


The gas turbine engines powering the most passenger and military aircrafts are also called jet engines. Tip

essor causes the leakage of high pressure fluid from pressure surface to


io, they are limited to short

However, the inability for spark ignition engines to work at high altitude warrants

further development to improve the efficiency of these miniature gas turbines. Hayami (1997) has found

s that axial movement of the casing has better efficiency over the movement of casing


erstand the flow behavior in order to




= 34.7 mm

o

and blade to blade view

S.M. Swamy, V. Panndurangadu and J.M. Shamkumar


Schematic diagram of centrifugal compressor setup

computational simulation of Centrifugal impeller with above specifications with 3mm thickness

throughout the blade for generation of three dimensional blades was done as sho

passage of the impeller with inlet at 50mm ahead of the impeller and outlet at a distance of 35mm

downstream of impeller is shown in

height. Three tip clearances of 2.2%, 5.1% and 7.9% of trailing edge blade height are obtained by moving

the casing axially. ANSYS CFX 15.0 version software is used for obtaining the solution and standard k

turbulence model is used for the closure. The computational domain is kept at

of the impeller in order to ensure that the inlet boundary conditions are not affected by the back pressure of

impeller blade. The blade gen layout with construction features related to wrapper angle, number of blades,

angles distribution and blade thickness for generation of three dimensional blades was done in order to

carry out structural analysis of the blade surfaces the software provided export options to generate

computational mesh accounting blade thickness is also explo

blade profiles generated in the software

block structured computational mesh software module called turbo grid. After these operations three

dimensional computational mesh generation process for compressor blade with boundary patches, periodic

surfaces are created.

Figure

Figure 3 Computational domain of single


IJMET/index.asp 180

of centrifugal compressor setup as shown in fig.1.The geometric model used for the


throughout the blade for generation of three dimensional blades was done as sho


downstream of impeller is shown in Fig. 3. Casing is designed with 0.7mm clearance throughout the blade

2.2%, 5.1% and 7.9% of trailing edge blade height are obtained by moving

the casing axially. ANSYS CFX 15.0 version software is used for obtaining the solution and standard k

turbulence model is used for the closure. The computational domain is kept at 50 mm ahead of the eye inlet


The blade gen layout with construction features related to wrapper angle, number of blades,

istribution and blade thickness for generation of three dimensional blades was done in order to


computational mesh accounting blade thickness is also explored in static structural analysis package. The

software mentioned above module are imported to template based multi



ure 2 Centrifugal compressor (without and with PS)

Computational domain of single passage (with and with


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The geometric model used for the


throughout the blade for generation of three dimensional blades was done as shown in Fig. 2. A single


Casing is designed with 0.7mm clearance throughout the blade

2.2%, 5.1% and 7.9% of trailing edge blade height are obtained by moving

the casing axially. ANSYS CFX 15.0 version software is used for obtaining the solution and standard k-ε

50 mm ahead of the eye inlet


The blade gen layout with construction features related to wrapper angle, number of blades,

istribution and blade thickness for generation of three dimensional blades was done in order to


structural analysis package. The

module are imported to template based multi-



Centrifugal compressor (without and with PS)

with and without PS)



3. COMPUTATIONAL METHOD

Most fluid flow simulation technology uses

compared to CFD simulation. And this is

the user is connecting up the dialogs within the

automatically, simply by dragging a connection between

of another. The geometry from either of the CAD

simulation model before setup option is activated. The next step is to generate the meshing solid

which several methods are provided in ANSYS

highly automated along with having a moderate to high degree of

meshing application meshing options panel displays, which allows you to

meshing preferences based on the phys

for this problem statement, with respective to this stationary, rotating components are defined. The density

going to vary compressible fluid with pressure fluid properties kept as air ideal gas. Reference pressure i

atm. Heat transfer is calculated by total energy because increase in density causes to increase in

temperature and because of rotational energy it will try to change the heat transfer rate, in total energy

methods only all this effects will get conside

because of near wall treatment is required and presence of ad

boundaries are pressure inlet with 0 Pa gauge pressure and mass flow 0.087 Kg/s at outlet respect

Rotating domain interface used as Stage. The stage averaging at the frame change interface incurs a one

time mixing loss. This loss is equivalent to assuming that the physical mixing supplied by the relative

motion between components is sufficiently

entering the downstream machine component. Stage analysis is most appropriate when the circumferential

variation of the flow is of the order of the component pitch stage for interfaces used

simulations. Physical Time scaling value was observed to be taken as 10

solution was decreased to 10-5

for better accuracy.

.


IJMET/index.asp 181

COMPUTATIONAL METHODOLOGY

Most fluid flow simulation technology uses completely different meshing, loading and solving

simulation. And this is where the true power of Workbench 1

the user is connecting up the dialogs within the schematic layouts, many of these are handled

automatically, simply by dragging a connection between the outputs of one analysis system and the inputs

another. The geometry from either of the CAD Systems and Blade Gen Model can be connected to

simulation model before setup option is activated. The next step is to generate the meshing solid

which several methods are provided in ANSYS Workbench 15.0 these tools have the benefit of being

automated along with having a moderate to high degree of user control. Upon startup of the

meshing options panel displays, which allows you to quickly and easily set your

the physics you are preparing to solve. In CFX we have to use turbo mode


going to vary compressible fluid with pressure fluid properties kept as air ideal gas. Reference pressure i



methods only all this effects will get considered. Shear Stress Transport Turbulence model was used

because of near wall treatment is required and presence of ad-versed pressure gradient. Inlet and outlet

boundaries are pressure inlet with 0 Pa gauge pressure and mass flow 0.087 Kg/s at outlet respect

Rotating domain interface used as Stage. The stage averaging at the frame change interface incurs a one


motion between components is sufficiently large to cause any upstream velocity profile to mix out prior to


variation of the flow is of the order of the component pitch stage for interfaces used

simulations. Physical Time scaling value was observed to be taken as 10-3

s. Accuracy criteria of RMS

for better accuracy.

Table 1 Mesh statistics

Frequency of distribution of skewness


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completely different meshing, loading and solving methods

where the true power of Workbench 15.0 comes into play. When

schematic layouts, many of these are handled

the outputs of one analysis system and the inputs

Systems and Blade Gen Model can be connected to

simulation model before setup option is activated. The next step is to generate the meshing solid body for

these tools have the benefit of being

user control. Upon startup of the

quickly and easily set your

In CFX we have to use turbo mode


going to vary compressible fluid with pressure fluid properties kept as air ideal gas. Reference pressure is 1



red. Shear Stress Transport Turbulence model was used

versed pressure gradient. Inlet and outlet

boundaries are pressure inlet with 0 Pa gauge pressure and mass flow 0.087 Kg/s at outlet respectively.

Rotating domain interface used as Stage. The stage averaging at the frame change interface incurs a one-


large to cause any upstream velocity profile to mix out prior to


variation of the flow is of the order of the component pitch stage for interfaces used for steady state

s. Accuracy criteria of RMS



Figure 4 Three dimensional mesh for single compressible blade row and streamlines from the inlet of impellor

around the blades

3.1. Validations for Tip Clearance

Prior to comparing the performance of the test cases, the numerical result for the three values of tip

clearances (τ=2.2%, 5.1% and 7.9%), was validated with the experimental data. Measurements were

carried out with steady probes and scanning box (Model FC091-3) and micro manometer (model FCO12).

Experimental results were validated by comparing the computed characteristic curves of the efficiency and

energy coefficient with the experimental data, static pressure from inlet to outlet of the blade with the

computational data. The CFD results showed satisfactory agreement with the experimental results over the

full range of operating conditions. The overall flow structure predicted by the numerical simulations,

including the jet-wake flow pattern, was qualitatively well captured, as observed by comparison with the

experimental results. Improved results may be obtained with PS.

4. RESULTS AND DISCUSSIONS

Performance Characteristics: Effect of Configuration: The effect of the two configurations on the

performance of the centrifugal compressor in Figure 5 shows the performance characteristic in terms of

energy coefficient, ψ vs. flow coefficient, φ for the three values of tip clearance for both configurations. At

all values of tip clearance, PS gives higher value of energy coefficient, as the partial shroud obstructs the

tip leakage flow and allows it to mix with main flow at a further distance. It may be observed that the

performance of the rotor without PS at τ=2.2% is almost similar to that the rotor with PS at τ=5.1%. Hence

by using partial shroud the rotor can operate at higher clearances with mechanical safe margins but without

sacrificing performance. Figure 6 shows the performance characteristic in terms of efficiency, η vs. flow

coefficient, φ for the three values of tip clearance for both configurations. At all values of tip clearance, PS

gives higher value of efficiency compared to that without PS due to reduced tip leakage flows. In fact the

efficiency of the rotor with PS even at the highest value of tip clearance observed is higher than the

efficiency of the rotor without partial shroud at the lowest value of tip clearance.

Effect of Tip Clearance: Effect of tip clearance of a low speed centrifugal compressor computational

investigation results can be concluded that partial shroud is very beneficial to the performance of the

compressor. In order to determine the effect of tip clearance on the performance of the compressor with

and without partial shroud means, the performance curves in terms of ψ vs. φ and η vs. φ are plotted for the

two configurations, viz., without PS, and with PS only for the three values of tip clearances observed, i.e. τ

= 2.2%, 5.1% and 7.9%. The ψ vs. φ curve shows that the operating range of the compressor is increased,

for the basic configuration, as the tip clearance is increased. The result of the numerical investigations

revealed that, energy coefficient across the compressor is reduced as the tip clearance is increased.



Figure 5 Performance of the centrifugal compressor: ψ vs. φ curve

Figure 6 Performance of the centrifugal compressor: η vs. φ

4.1. Velocity Vectors at Outlet

A velocity vector at the exit of the impeller is shown in figure 7. The fluid flows with PS on tip of the

blade velocities are high on suction surface than pressure surface because of blade curvature. With increase

in tip clearance, the velocity on both pressure side and suction side is decreasing. For τ = 2.2%, 5.1% and

7.9% clearance for without PS. Configuration with PS, high velocity of the fluid above the blade from

pressure side to suction side through tip clearance is observed. The velocity distribution at outlet is

minimum near hub and casing indicating the boundary layer, velocity is observed near the casing,

indicating higher energy transfer near casing. Velocity for 5.1%, 7.9% tip clearance have more velocity at

the outlet than 2.2% tip clearance with partial shroud indicating loss of static pressure at outlet.

0

0.5

1

1.5

2

2.5

0 0.1 0.2 0.3 0.4

En

erg

y C

oe

ffic

ien

t, ψ

Flow Coefficient, φ

CFD_2.2%_With_PS

CFD_2.2%_Without_PS

CFD_5.1%_With_PS

CFD_5.1%_Without_PS

CFD_7.9%_With_PS

Exp._2.2%_Without_PS

Exp._2.2%_With_PS

Exp._5.1%_With_PS


Exp._7.9%_With_PS


0

10

20

30

40

50

60

0 0.1 0.2 0.3 0.4 0.5

Eff

icie

ncy

, (η

)

Flow Coefficient (φ)

2.2%_CFD_With_PS

2.2%_CFD_Without_PS

5.1%_CFD_With_PS

5.1%_CFD_Without_PS

7.9%_CFD_With_PS

7.9%_CFD_Without_PS

2.2%_Exp._With_PS

2.2%_Exp_Without_PS

5.1%_Exp._With_PS

5.1%_Exp._Without_PS

7.9%_Exp._With_PS



Velocity vectors at exit of the blade



Figure 7 Velocity vectors at exit of the blade

4.2. Static Pressure Contours at Span 0.7

Static pressure contours in blade to blade view, at span 0.7

gradual pressure rise from inlet to outlet of the compressor due to dynamic action of the rotating impeller.

Gradual increase of static pressure from inlet to outlet is clearly observed at all tip clearances

5.1% and 7.9% clearance. With PS

tip of the blade, the pressure at outlet is reduced. High pressure on pressure side of the blade and low

pressure on suction side of the blade are observed at all tip clearances. With

reduction in pressure on both pressure side and suction side is found.


IJMET/index.asp 184

Velocity vectors at exit of the blade for 2.2% clearance for without PS and with PS

Velocity vectors at exit of the blade for 5.1% clearance for without PS and with PS

Velocity vectors at exit of the blade for 7.9% clearance for without PS and with

Velocity vectors at exit of the blade for τ = 2.2%, 5.1% and 7.9% clearance for without PS and with PS

Static Pressure Contours at Span 0.7

ressure contours in blade to blade view, at span 0.7 are shown in figure


Gradual increase of static pressure from inlet to outlet is clearly observed at all tip clearances

. With PS on tip of the blade, low pressure change is observed. But without PS on


pressure on suction side of the blade are observed at all tip clearances. With

reduction in pressure on both pressure side and suction side is found.


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for 2.2% clearance for without PS and with PS

% clearance for without PS and with PS

for 7.9% clearance for without PS and with PS

clearance for without PS and with PS

figure 8. The contours show


Gradual increase of static pressure from inlet to outlet is clearly observed at all tip clearances i.e. τ = 2.2%,

on tip of the blade, low pressure change is observed. But without PS on


pressure on suction side of the blade are observed at all tip clearances. With increase in tip clearance,



Static pressure contours for 2.2% clearance at mid span without PS and with PS

Static pressure contours for 5.1% clearance at mid span without PS and

Static pressure contours for 7.9% clearance at span

Figure 8 Static pressure contours for

4.3. Velocity Contours at Span 0 .7

Velocity contours in blade to blade view, at span 0.7 is shown in

improved velocities on suction side with PS.

reducing and also velocity improvement is observed.

side with partial shroud. The low velocity passage wake area on suction side of the blade is reducing with

PS on tip of the blade. At 2.2% clearance, velocities are more near the suction side shroud corner a

relative velocities are low near pressure side hub corner. Due to the centrifugal forces and curvature,

velocity on suction side is higher than pressure side. For other clearances, leakage of flow from pressure

side to suction side of the blade through

blade meets the tip leakage flow near suction side shroud corner.


IJMET/index.asp 185

pressure contours for 2.2% clearance at mid span without PS and with PS

pressure contours for 5.1% clearance at mid span without PS and

pressure contours for 7.9% clearance at span 0.7 without PS and with PS

pressure contours for τ = 2.2%, 5.1% and 7.9% clearance at mid span without PS and with PS

Velocity Contours at Span 0 .7

Velocity contours in blade to blade view, at span 0.7 is shown in figure 9. The velocity contours show

improved velocities on suction side with PS. With PS on tip of the blade, the low velocity region is

reducing and also velocity improvement is observed. The contours show improved velocities on suction


At 2.2% clearance, velocities are more near the suction side shroud corner a



side to suction side of the blade through the tip gap is observed. Vertical flow near the suction side of the

blade meets the tip leakage flow near suction side shroud corner. Outlet to inlet pressure ratio at different


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without PS and with PS

= 2.2%, 5.1% and 7.9% clearance at mid span without PS and with PS

The velocity contours show

With PS on tip of the blade, the low velocity region is

The contours show improved velocities on suction


At 2.2% clearance, velocities are more near the suction side shroud corner and



the tip gap is observed. Vertical flow near the suction side of the

Outlet to inlet pressure ratio at different



flow coefficients is increase in pressure ratio with PS on tip of the blade is

coefficients. But without PS on tip of the blade, reduction in pressure ratio is observed at all flow

coefficients because of more fluid leakage through the basic configuration.

Velocity contours for 2.2% clearance for without PS a

Velocity contours for 5.1% clearance for without PS and with PS

Velocity contours for

Figure 9 Velocity contours for τ

Blade loading: Blade loading ch

shown in fig. 11. Fluid flow without PS and with PS on tip of the blade, significant change in pressure on

suction side is observed. Low static pressure on

observed. With increase in tip clearance, static pressure on both pressure side and suction side are

reducing.


IJMET/index.asp 186

flow coefficients is increase in pressure ratio with PS on tip of the blade is


coefficients because of more fluid leakage through the basic configuration.

contours for 2.2% clearance for without PS and with PS


contours for7.9% clearance for without PS and with PS

τ = 2.2%, 5.1% and 7.9% clearance at span 0.7 for without PS and with PS

Blade loading charts at design for 2.2% tip clearance at without PS and with PS


suction side is observed. Low static pressure on suction and high pressure on pressure side of the blade is



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flow coefficients is increase in pressure ratio with PS on tip of the blade is observed at all flow


nd with PS


% clearance for without PS and with PS

for without PS and with PS

for 2.2% tip clearance at without PS and with PS were


suction and high pressure on pressure side of the blade is




Figure 10 Blade loading for 2.2% clearance at without and with partial shroud

Static Pressure Distribution: Static pressure distribution along stream

is shown in figure 12. This plot indicated

increasing gradually along the stream wise direction due to the dynamic head

impeller. A drop in static pressure near stream

to the acceleration of the flow in to the eye of the impeller. Static pressure reduction is observed with PS

on tip of the blade. With PS on tip of the blade, static pressure at outlet is more than the static pressure

basic configuration.

Figure 11 Pressure inlet to outlet over stream

5. CONCLUSIONS

Tip clearance effects on flow field of a low speed centrifugal compressor

(PS) attached to the rotor blade tip at three values of tip clearance, viz.

blade height at the exit at four flow coefficients, namely,

φ=0.28 (design flow coefficient) and

in the tip clearance region. Effect of tip clearance of a low speed centrifugal compressor computational

investigation results can be concluded that par

compressor, for both the configurations. In order to determine the effect of tip clearance on the

performance of the compressor with and without passive means, the performance curves in terms of

φ and η vs. φ are plotted for the two configurations, viz., without PS, and with PS only for the three values

of tip clearances observed, i.e. τ = 2.2%, 5.1% and 7.9%.

at span 0.7 is the gradual pressure rise from inlet to outlet of the compressor due to dynamic action of the

rotating impeller. The velocity contours show improved velocities on suction side with PS. With PS

of the blade, the low velocity region is reducing and also velocity improvement is observed

100500

101000

101500

102000

102500

103000

0

Sta

tic

Pre

ssu

re,

(Pa

)


IJMET/index.asp 187

Blade loading for 2.2% clearance at without and with partial shroud

Static pressure distribution along stream wise direction from inlet to outlet

indicated that the pressure from inlet to the outlet of the compressor is

increasing gradually along the stream wise direction due to the dynamic head

impeller. A drop in static pressure near stream wise direction of 0.2 is observed


on tip of the blade. With PS on tip of the blade, static pressure at outlet is more than the static pressure

Pressure inlet to outlet over stream wise for 2.2% clearance without

Tip clearance effects on flow field of a low speed centrifugal compressor without and with partial

(PS) attached to the rotor blade tip at three values of tip clearance, viz. τ=2.2%, 5.1% and 7.9% of rotor

flow coefficients, namely, φ=0.12, 0.18 (below design flow coefficient),

=0.28 (design flow coefficient) and φ=0.34 (above design flow coefficient),

Effect of tip clearance of a low speed centrifugal compressor computational

investigation results can be concluded that partial shroud is very beneficial to the performance of the


performance of the compressor with and without passive means, the performance curves in terms of

are plotted for the two configurations, viz., without PS, and with PS only for the three values

= 2.2%, 5.1% and 7.9%. Static pressure contours in blade to blade view,


The velocity contours show improved velocities on suction side with PS. With PS

of the blade, the low velocity region is reducing and also velocity improvement is observed

1 2 3

Stream Wise Location

2.2%_With_PS

2.2%_Without_PS


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Blade loading for 2.2% clearance at without and with partial shroud

wise direction from inlet to outlet

that the pressure from inlet to the outlet of the compressor is

increasing gradually along the stream wise direction due to the dynamic head developed by the rotating

wise direction of 0.2 is observed at 2.2% tip clearance due


on tip of the blade. With PS on tip of the blade, static pressure at outlet is more than the static pressure at

without and with PS

without and with partial shroud

=2.2%, 5.1% and 7.9% of rotor

(below design flow coefficient),

=0.34 (above design flow coefficient), is analyzed computationally

Effect of tip clearance of a low speed centrifugal compressor computational

tial shroud is very beneficial to the performance of the


performance of the compressor with and without passive means, the performance curves in terms of ψ vs.

are plotted for the two configurations, viz., without PS, and with PS only for the three values

Static pressure contours in blade to blade view,


The velocity contours show improved velocities on suction side with PS. With PS on tip

of the blade, the low velocity region is reducing and also velocity improvement is observed.

2.2%_With_PS

2.2%_Without_PS



REFERENCES

[1] Fowler, H. S., (1971) Aerodynamics Tests on a Centrifugal Fan Impeller Model with Swept Back

Blades, N.R.C. Report No. 12020.

[2] Senoo, Y. And M. Ishida, (1975) Behaviour of Severely Asymmetric Flow in a Vaneless Diffuser,

Trans. ASME, Jl. of Engg. for Power, 97, 375-382

[3] Moore, J., (1973), "A Wake and an Eddy in a Rotating Radial-Flow Passage", Part 1: Experimental

Observation", Part 2:"Flow Model", Trans. ASME Journal of Engineering for Power.pp.205-219.

[4] Eckardt, D., (1976) Detailed Flow Investigations within a High Speed Centrifugal Compressor Impeller,

Trans. ASME, Jl. of Fluids Engg. 98, 390-402.

[5] G. Sravan Kumar, Dr. D. Azad and K. Mohan Laxmi, Valuation of Turbulence Modelling on Low

Speed Centrifugal Compressor Using Computational Fluid Dynamics. International Journal of

Mechanical Engineering and Technology (IJMET), 7(6), 2016, pp. 634–641.

[6] Adler, D., (1980) Status of Centrifugal Internal Aerodynamics, Part II: Experiments and Influence of

Viscosity, Trans. ASME, Jl. of Engineering for Power, 102, 738-746

[7] Shivaprasad, B. G. and B. R. Ramaprian, (1977) Turbulence Measurements in Boundary Layers along

Mildly Curved Surfaces, ASME Paper 77-WA/FE-8.

[8] Krain, H. and D. Eckardt, (1978) The Flow Field in a High Speed Centrifugal Impeller-A Comparison

of Experimental and Theoretical Results, First International Conference on Centrifugal Compressor

Technology, IIT Madras.

[9] Johnson, M.W., and Moore, J., (1980), "The development of wake flow in a centrifugal compressor"

ASME Journal of Engineering for Power, Vol. 102, pp. 383-390.

[10] Johnson, M. W. and J. Moore, (1983) Secondary Flow Mixing Losses in a Centrifugal Impeller, Trans.

ASME, Jl. of Engg. for Power, 105, 24-32.

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